1. Field of the Invention
The present invention relates generally to a heat exchanger for transferring heat between a gas and a liquid and, more particularly, to a fin-type heat exchanger having porous fins on the gas side that are positioned within a gas flow chamber of the heat exchanger such that all, or substantially all, of the gas is forced to pass or transpire through the large number of pores in the fins to enhance heat transfer by increasing overall heat transfer surface area and heat transfer coefficient.
2. Description of Related Art
Heat exchangers are used extensively in industrial and consumer applications, and typically employ two moving fluids, one fluid being hotter than the other, to transfer heat to the colder fluid. Many heat exchangers currently in use, such as in air conditioners, automotive radiators, process industry air-cooled condensers, and boilers, transfer heat between a gas and a single or multi-phase liquid. Typically, such heat exchangers include a number of liquid conduits, e.g., circular, oval, or flat tubes, conduits defined by plates, and the like, that are positioned within a shell or casing which defines a gas flow passage or chamber. The heat exchanger uses a fan or blower to force a gas, e.g., air, to flow within the gas flow chamber in a perpendicular (i.e., cross-flow) or parallel (i.e., counter-flow) direction relative to the liquid conduits. The resulting heat transfer between the liquid and the gas is directly proportional to the heat transfer surface area between the liquid and the gas, the temperature difference between the liquid and the gas, and the overall heat transfer coefficient of the heat exchanger. The overall heat transfer coefficient is defined in terms of the total thermal resistance to heat transfer between the gas and the liquid, and it is dependent on a number of characteristics of the heat exchanger design, such as the thermal conductivity of the material used to fabricate the conduit and the local film coefficients along the conduit, i.e., measurements of how readily heat can be exchanged between the gas and the exterior surfaces of the conduit.
Although gas-liquid heat exchangers are widely used, the heat transfer per degree of temperature difference between the hot and cold sides of these heat exchangers is quite low due in large part to the low density and low thermal conductivities of gases. This heat transfer per degree of temperature difference can be stated mathematically as the product (UA) of the overall heat transfer coefficient (U) and the heat exchange area (A). Low UA leads to relatively high operating and capital costs for gas-liquid heat exchangers because a greater number of units and/or larger capacity units that require more power must be used to account for this low UA in obtaining a desired heat transfer. For example, geothermal power plants operate at low temperature differences between the gas and the liquid and, in these power plants, more than 25 percent of the cost of producing electricity is the expense of purchasing and operating gas-liquid heat exchangers, i.e., condensers. As a result of these high costs, continuing efforts are being made to improve the UA of gas-liquid heat exchangers while at the same time controlling the manufacturing and operating cost to increase the likelihood that new heat exchanger designs will be adopted by industry and consumers.
Finned-tube heat exchangers have been used for many years to improve the gas-side heat transfer rate by increasing the heat transfer surface area available for contacting the gas as it flows through the heat exchanger. In general, finned-tube heat exchangers are cross-flow heat exchangers that include a number of tubes, i.e., conduits, for carrying the liquid fabricated from aluminum, copper, steel, or other high thermal conductivity materials. The tubes pass through and contact a series of parallel, high thermal conductivity material sheets or plates, i.e., fins, which provide an extended heat transfer area for the tubes. The overall heat transfer area is based on the number and size of included fins, with the typical number of fins used ranging from five to fifteen fins per inch. The fins define parallel channels that direct the gas flow across and among the tubes. Heat transfer occurs as the gas flows along and contacts the surface of the fins and as the gas contacts the outer surfaces of the tubes. The highest heat transfer rate on a flat surface like a flat fin occurs at the leading edge of the surface and decreases with distance from the leading edge as a boundary layer develops and thickens causing the local heat transfer coefficient to decrease. However, although finned-tube heat exchangers are widely used because they are relatively inexpensive to produce and do not create a large pressure drop, there are several operational drawbacks to finned-tube heat exchangers. For example, finned-tube heat exchangers have low heat transfer coefficients on large portions of the fins due to the development of thick boundary layers. Additionally, these heat exchangers have poor heat transfer in the wake or shadowed regions behind tubes as a majority of the gas flowing over a tube does not contact the backside of the tube or contact the portion of the fin surface that is shadowed by the tube.
In an attempt to increase the effectiveness of finned-tube heat exchangers, efforts have been made to vary the surface and overall geometry of the parallel fins to interrupt gas boundary layers or to make it more difficult for thick boundary layers to form on the fins. For example, finned-tube heat exchangers have utilized triangular or s-shaped wavy fins to enhance the heat transfer coefficient by disrupting boundary layer development and, also, by increasing the available heat transfer area. Alternatively, the surface geometry of flat, parallel fins can be enhanced, as is often done in refrigerant condensers, by slitting the fin three or four times in the areas of the fin between the tubes, thereby interfering with boundary layer development by creating offset surfaces on the fin that cause repeated growth and wake destruction of boundary layers. Another fin geometry sometimes used on the gas side of heat exchangers, but more often on the liquid side of heat exchangers such as automobile radiators, are accordion-like, louvered sheets that define parallel, triangular-shaped channels through which the gas flows. The formation of boundary layers is disrupted by the shape of the louvered-surface as the majority of the gas flows along the fin in the channel and also by the flow of a small amount of the gas through the louvers into adjacent channels.
U.S. Pat. No. 4,768,563 issued to Tsukamoto et al. discloses a finned-tube heat exchanger with corrugated and perforated fins that are arranged on staggered tubes so as to define parallel fluid channels across the tubes. The corrugated fins are positioned ridge to ridge and valley to valley so that the fluid channels have alternating expanding and contracting flow sections. This fin arrangement establishes differences in fluid pressures in the gas in adjacent fluid channels because expanding flow sections are positioned adjacent contracting flow sections. With this fin configuration, the main gas flow is along the parallel fluid channels, and boundary layer development on the fins is at least partially disrupted by the corrugated surfaces of the fins. Additionally, a small secondary flow is developed between adjacent fluid channels due to the differences in the fluid pressures in adjacent fluid channels causing a small portion of the gas to breathe or flow through the perforations into adjacent fluid channels and further disrupt the boundary layers.
U.S. Pat. No. 3,804,159 issued to Searight et al. discloses a pleated fin and tube cooling coil that attempts to use well-known jet impingement technology to enhance heat transfer on the back side of the fins. According to Searight et al. jet impingement on the back sides of the cooling fins is obtainable by forcing cooling gas through a small number of perforations in the fins at relatively high velocity to contact the backside of the adjacent fin. In this regard, Searight et al. uses low-porosity fins, i.e., less than 20 percent and preferably between 2 and 15 percent open fin area, to obtain high jet speeds when a large volume of gas is forced through a small number of holes and uses tightly spaced fins, i.e., 12 fins per inch, to allow the jets of gas to reach the adjacent fin. Additionally, the holes are relatively large in diameter, typically much greater than the thickness of the fin, to increase the jet size. Jet impingement requires careful staggering of the perforations on each adjacent fin so that jets strike adjacent fins between the perforations. While potentially increasing heat transfer on only the back sides of the pleated fins, the disclosed fin arrangement and design results in serious problems with high pressure drops caused by the close fin spacing and the low porosity of the fins. The resulting high pressure drop through the disclosed cooling coil significantly increases fan power requirements thereby lowering overall UA of the heat exchanger relative to a non-perforated, parallel fin heat exchanger.
While some of the above changes in the fin surface and fin shape may provide somewhat higher heat transfer coefficients in finned-tube heat exchangers, the UA of heat exchangers that include these enhanced fins remains relatively low. This low UA is, at least in part, due to ongoing problems with low heat transfer coefficients on the gas side and poor heat transfer in shadowed or wake regions behind the tubes. Further, many of the above design changes result in unacceptably large increases in pressure drop on the gas side of the heat exchanger that require increased expenditures on fan power.
Consequently, in spite of the well-developed state of heat transfer technology, there remains a need for a more effective gas-liquid heat exchanger that provides improved heat transfer capabilities while controlling operating and capital costs to make implementation cost effective for industrial and consumer applications.
Accordingly, it is a general object of the present invention to provide a gas-fluid heat exchanger with an increased UA value and improved ratio of UA to pressure drop.
It is a related object of the present invention to provide a gas-fluid heat exchanger with improved heat transfer properties on the gas side.
It is another related object of the present invention to provide a more effective gas-fluid heat exchanger that can be economically operated and manufactured with present technologies.
It is a more specific object of the present invention to provide a gas-fluid heat exchanger with an enhanced fin geometry and surface configuration that enhances heat transfer properties on the gas side of the heat exchanger.
Additional objects, advantages, and novel features of the invention are set forth in part in the description that follows and will become apparent to those skilled in the art upon examination of the following description and figures or may be learned by practicing the invention. Further, the objects and the advantages of the invention may be realized and attained by means of the instrumentalities and in combinations particularly pointed out in the appended claims.
To achieve the foregoing and other objects and in accordance with the purposes of the present invention, as embodied and broadly described herein, one preferred embodiment of the invention includes a fin and tube assembly for positioning in a gas flow path of a gas-fluid heat exchanger such that a gas is forced to flow through the fin and tube assembly to significantly increase the heat transfer surface area available (A) and the heat transfer coefficient (U) while also controlling any corresponding pressure drop. The fin and tube assembly includes fluid conduits, which are tubes in one embodiment, for directing a fluid through the heat exchanger and a heat transfer element in heat conductive contact with the fluid tubes to provide an extended heat transfer area between the fluid in the fluid conduits and the flowing gas. The heat transfer element is corrugated to have a cross-sectional shape of ridges and grooves with a heat transfer fin formed between each ridge and groove. The heat transfer element is positioned transverse to the gas flow path such that gas is forced to flow down along the fins from the ridges to the grooves in a flow channel.
The fins are highly porous, e.g., perforated, sintered, expanded, stabbed, built up from layers, and the like, to contain numerous orifices or pores to provide flow passages for the gas to transpire through the fins. Each of the interior surfaces of the orifices or pores contribute to the overall heat transfer surface area of the fin and tube assembly, which results in a significant increase in heat transfer surface area and a corresponding increase in the heat transfer rate of the fin and tube assembly. The grooves are closed or only partially open so that all or a substantial portion of the gas transpires through the orifices in the fins to provide an increased amount of heat transfer area including the interior surfaces of the orifices as well as the front and back surfaces of the fins and to establish desirable gas flow patterns that control boundary layer development and otherwise increase heat transfer rates within the fin and tube assembly. In a preferred embodiment, the fins are highly porous, i.e., 25 percent or considerably higher such as 50 to 70 percent or higher, with small sized holes, e.g., a diameter for a round hole of about the thickness of the fin, such that the interior surface area provides a large increase over nonporous fins, e.g., about a 25 percent or larger increase for higher porosities, in the available heat transfer surface area of the heat transfer element.